Apparatus for controlling the characteristics of fluid pressure operated friction type power absorption devices

ABSTRACT

Apparatus for controlling a rotary power absorber while absorbing driving torque produced by a prime mover. The power absorber includes a housing containing fluid pressure operated brake elements that are actuatable hydraulically to provide retarding force. The control means for the brake applying means includes a governor-controlled air valve and a transducer which converts air pressure into hydraulic pressure for actuating the brake elements. The governor is driven at a speed proportional to the speed of rotation of the prime mover and actuates the air valve so that fluid pressure is applied to the brake elements in direct proportion to the speed of the prime mover. An adjustable linkage is interposed between the governor and the air valve for controlling the maximum retarding force at various engine speeds. The linkage includes two parallel bars pivoted at opposite ends, with the governor acting on the free end of one bar and the free end of the other bar acting on the air valve. A roller is positionable at any desired point along the length of the bars to vary the ratio of air valve movement to governor movement, as desired, to correspondingly vary the retarding force. The retarding force is increased and decreased at a rate faster than the changes in speed and torque of the prime mover to avoid stalling of the prime mover as frequently occurs when a constant load is sought to be applied to the prime mover and there is a momentary failure in power of the prime mover.

United States Patent Cline [72] Inventor: Edwin L. Cline, Pasadena,Calif. [73] Assignee: Clayton Manufacturing Company,

El Monte, Calif.

[22] Filed: July 3, 1969 [21] Appl. No.: 839,005

Related U.S. Application Data [62] Division of Ser. No. 559,490, July22, 1966,

Pat. No. 3,453,874.

[52] U.S. Cl ..73/135, 73/117 [51] Int. Cl. ..G0ll 3/16 [58] Field ofSearch ..73/117, 134, 135; 318/304, 318/372; 188/180, 182

[56] References Cited UNITED STATES PATENTS 1,726,599 9/1929 Wasson..l'88/180 2,012,110 8/1935 Shroyer ..73/135 X 2,220,007 10/1940 Wintheret al ..73/134 2,266,213 12/1941 Kattwinkel ..188/181 X 3,050,993 8/1962Draughon et a1. ..73/134 3,068,689 12/1962 Warsaw ..73/135 3,193,0577/1965 Rudqvist et al ..188/182X FOREIGN PATENTS OR APPLICATIONS 279,9147/1929 Great Britain ..73/135 51 Oct. 17, 1972 Primary Examiner--CharlesA. Ruehl Attorney-Bacon & Thomas [57] ABSTRACT Apparatus for controllinga rotary power absorber while absorbing driving torque produced by aprime mover. The power absorber includes a housing containing fluidpressure operated brake elements that are actuatable hydraulically toprovide retarding force, The control means for the brake applying meansincludes a governor-controlled air valve and a transducer which convertsair pressure into hydraulic pressure for actuating the brake elements.The governor is driven at a speed proportional to the speed of rotationof the prime mover and actuates the air valve so that fluid pressure isapplied to the brake elements in direct proportion to the speed of theprime mover. An adjustable linkage is interposed between the governorand the air valve for controlling the maximum retarding force at variousengine speeds. The linkage includes two parallel bars pivoted atopposite ends, with the governor acting on the free end of one bar andthe free end of the other bar acting on the air valve. A roller ispositionable at any desired point along the length of the bars to varythe ratio of air valve movement to governor movement, as desired, tocorrespondingly vary the retarding force. The retarding force isincreased and decreased at a rate faster than the changes in speed andtorque of the prime mover to avoid stalling of the prime mover asfrequently occurs when a constant load is sought to be applied to theprime mover and there is a momentary failure in power of the primemover.

12 Claims, 7 Drawing Figures PATENIEDUCT 11 m2 3. 698,243

7 sum 1 or 5 Fowl/v L. CL/NE H TTOPNEYS PATENTEDucr 17 m2 SHEET 2 OF 5ATTORNEYS APPARATUS FOR CONTROLLING THE CHARACTERISTICS OF FLUIDPRESSURE OPERATED FRICTION TYPE POWER ABSORPTION DEVICES CROSS-REFERENCEThis application is a division of my co-pending application Ser. No.559,490 filed July 22, 1966 now US. Pat. No. 3,453,874.

BACKGROUND OF THE INVENTION 1. Field of the Invention This inventionrelates to equipment for testing the under-load performancecharacteristics of a prime mover by rotary friction type powerabsorption devices, and more particularly to novel load control meanstherefor. For convenience, such power absorption device will hereinaftersimply be referred to as a friction absorber.

2. Description of the Prior Art Currently available friction absorberscomprise a rotating brake drum or disk to be connected with the outputshaft of the prime mover, and stationary friction pads or brake shoesthat are engageable with the drum or disk to apply a retarding forcethereto by frictional contact. The degree of retarding action isdictated by the force with which this frictional contact is made. Thesystem that is used to apply this force is referred to as the LoadControl System. In simple friction absorbers this is a fixed force andonly one speed versus power curve is possible for the reason that aconstant retarding force is applied regardless of the speed or torquethat is developed by the engine. In more flexible friction absorbers,the load control will allow the operator to manually vary the force offrictional contact. In the case of a hydraulically actuated system, forexample, this is accomplished by the operator varying the fluidpressure. A series of speed versus power curves can then be obtained.

Due to the speed versus power characteristics of friction absorbers, thefixed force load control system is unsatisfactory because, for a givencontact pressure of the brake shoes, the retarding force remainsconstant (neglecting the effects of temperature) and the same amount oftorque will be absorbed throughout the speed range. Friction absorbersthat produce a constant retarding force are further objectionable, inthat such devices will cause the prime mover to stall in the event thatthe engine should momentarily miss or lose power for any reason. Suchconstantforce friction absorbers are still further objectionable fromthe standpoint that they are unstable over the range in which theretarding force is equal to engine torque, which may cover a substantialspeed range.

Since horsepower involves both speed and torque, the horsepower willincrease and decrease directly with speed. In testing engines withfriction absorbers, speed stability can be acquired only when theretarding force of the friction absorber increases and decreases withspeed faster than that of the prime mover being tested. An idealcondition would be for the friction retarding force to start at zerowith zero speed and change as a square of the speed change. Thisrelationship is very close to the load imposed on a conventionalautomobile engine when the vehicle is driven on a level road and, hence,represents highly desirable loading characteristics to be simulated inpractice. Such operational characteristics obviously cannot be attainedwith the fixed force load control for reasons stated above. Likewise, itis extremely difficult and practically impossible to establish andmaintain such operating characteristics in a friction absorber by manualcontrol of the force load, and, hence, such manual control leaves muchto be desired.

SUMMARY OF THE INVENTION Accordingly, there has long existed the needfor load control means for friction absorbers that will render the samepractical and avoid the principal objections thereto noted above. Theload control means of the present invention is designed to obviate theseobjections and may take any number of forms. Each load control means ismade to simulate road conditions, to maintain stability at any speed,and to control the friction absorber means so that the retarding forceof the friction absorber means is zero at zero engine speed and willrise and fall faster than the torque of the prime mover being tested.The present load control means contemplates a fluid control system thatis responsive to engine speed, and which can be adjusted and the loadpre-selected to impose a retarding force of a given value at a givenspeed and automatically and correspondingly control the operatingcharacteristics of the friction absorber in accordance with theforegoing relationship at all other speeds. The system is also capableof being remotely controlled by an operator.

The present load control means is applicable in principle to all typesof friction absorbers, irrespective of whether the friction absorber isdirectly or indirectly coupled with the output shaft of the prime mover.A direct method would be to connect the prime mover shaft directly tothe input shaft of the friction absorber, as in an engine dynamometertest. stand setup. An indirect method would involve the incorporation ofthe friction absorber in a chassis dynamometer for testing engines ofautomobiles or trucks without removing the engine from the vehicle. Insuch case, rolls are usually provided to form a type of treadmill forthe drive wheels of the vehicle and the friction absorber is thenconnected with a driven roll. Power from the engine would then benormally transmitted to the friction absorber through the vehicletransmission and differential. For illustrative purposes, and not by wayof limitation, the load control means of the present invention is shownand described in connection with a friction absorber associated with achassis dynamometer.

More specifically, the invention comprises a friction power absorberincluding a housing containing fluid pressure operated friction brakeelements that are actuatable to provide retarding force. The controlmeans for the brake applying means includes a governor driven at a speedproportional to the speed of rotation of the prime mover so that linearmovement of an element of the governor is produced in direct proportionto the speed of the prime mover. Such movement is transmitted through anadjustable linkage system to an air pressure control valve, whichactivates a transducer to apply hydraulic pressure to the brake elementsfor controlling the maximum brake retarding force at various enginespeeds. The retarding force is increased and decreased at a rate fasterthan the changes in speed and torque of the prime mover. The brakeactuating pressure can be varied as the square of the speed of the primemover, or any other suitable mathematical function of the speed of theprime mover.

Accordingly, the principal object of the present invention is to providea load control apparatus for controlling the retarding forcecharacteristics of a friction absorber, so that the retarding forceproduced thereby increases and decreases in value at a rate faster thanthe increases and decreases in driving torque applied to said absorberfrom a prime mover, and so that the applied retarding force value forany given prime mover speed can be pre-selected at will.

Another object is to provide a dynamometer, including a rotary frictionabsorber and control means for creating retarding force in opposition todriving torque applied thereto from a prime mover, designed so that thevalue of the retarding force will be changed in proportion to theoccurring changes in driving torque and speed applied to said frictionabsorber by the prime mover.

A further object is to provide adjustable control means for a frictionabsorber, constructed so that various value relationships betweenretarding force and driving torque can be pre-selected.

Another object is to provide apparatus for controlling the frictioncharacteristics of a friction absorber constructed to automatically varythe retarding force produced by the friction absorber in a preselectedmanner and as a function of the speed of rotation of the shaft supplyingdriving torque to said unit.

A more specific object is to provide load control means for a frictionabsorber, wherein means responsive to the speed of the prime mover beingtested is utilized to regulate a fluid pressure control system toactuate a power absorber in accordance with speed changes of the primemover, to correspondingly vary the retarding force created by saidfriction absorber to provide a pre-selected load on the prime mover at agiven speed.

Another object is to provide load control means for a friction absorberthat can be remotely controlled and which allows pre-selection of thedegree of load to be applied to a prime mover.

Other objects and many of the attendant advantages of the invention willbecome readily apparent from the following description, when takentogether with the accompanying drawings.

DESCRIPTION OF THE DRAWINGS FIG. 1 is a fragmentary, diagrammatic planview of a chassis dynamometer incorporating a friction absorbercontrolled by a pneumatic-hydraulic pressure transducer, an air controlvalve for supplying air under pressure to said transducer, and agovernor or centrifugal force-responsive means for actuating said aircontrol valve;

FIG. 2 is a diagrammatic view of the control system of FIG. 1, showingin cross-section the pneumatichydraulic transducer, the air pressurecontrol valve therefor, and the centrifugal force-responsive means foractuating the air pressure control valve;

FIG. 3 is an enlarged fragmentary sectional view, taken along the line33 in FIG. 2;

FIG. 4 is an end elevational view of the friction absorber as seen alongthe line 4-4 in FIG. 1, and showing a torque arm for actuating apressure transmitting device connected to a gauge for indicating thetorque being absorbed;

FIG. 5 is an enlarged vertical sectional view through the frictionabsorber, taken along the line 5-5 in FIG. 4,

showing certain details of construction of the friction absorber, andthe manner in which it is mounted in a coolant casing;

FIG. 6 is a graph comparing engine driving torque in foot pounds andfriction absorber retarding force, with vehicle speed in miles per hourand showing in particular by the curves retarding force produced bycontrolling differently the pressure applied to the friction elements;and

FIG. 7 is a graph comparing road horsepower absorbed by the frictionabsorber with vehicle speed in miles per hour, and showing typical powercurves for the friction absorber resulting from using the load controlsystems of the present invention.

DESCRIPTION OF A PREFERRED EMBODIMENT Referring now to FIG. 1, a portionof a conventional chassis dynamometer for testing motor vehicles isshown for use in conducting under-load testing of the engine of themotor vehicle, the dynamometer including absorber assembly 2 to whichdriving torque is supplied by an input shaft 4 supported by spacedbearings 6. Driving torque is transmitted indirectly from the engine ofthe motor vehicle undergoing test to the input shaft 4 by a rollassembly 8, upon which the driving wheels (not shown) of the motorvehicle are supported.

The roll assembly 8 has a generally rectangular frame comprised oflongitudinal side members 10, interconnected by transverse end members12, only one of which is shown. The dynamometer includes two pairs ofrolls, one of which pairs consists of the parallel rolls 18 and 20mounted on shafts 22 and 24, respectively, which are supported bybearings 26 mounted on the transverse members 12. The shaft 22 extendsbeyond the end member 12, and is connected by a coupling 40 with theinput shaft 4 of the power absorber assembly 2. Thus, when the engine ofa motor vehicle positioned with one of its drive wheels disposed on therolls 18 and 20 is operated to rotate said rolls, the wheel will drivethe roll 18 to thereby transmit driving torque from the vehicle engineto the input shaft 4 of the power absorber assembly 2.

Referring now in particular to FIGS. 4 and 5, the power absorberassembly 2 includes a coolant-receiving casing 42. The front of thecasing 42 is open and is surrounded by an external flange 52. A coverplate 54 and a gasket 55 are mounted on the flange 52 to close and sealthe casing 42, and are secured to said flange by bolts 56. The casing 42has an inlet pipe 57 connecting to an opening in the bottom wall 48 foradmitting liquid coolant 58 into the casing. An outlet pipe 59 isconnected to an opening near the top of the rear wall 44 of the casingfor the discharge of said coolant.

The cover plate 54 has a centrally positioned boss 72 in which the inputshaft 4 is rotatably mounted. The rear wall 44 of the casing 42 has acentral boss through which extends a bore 122 aligned with a bore 74 ina boss 72 on the cover plate 54.

Mounted within the casing 42 is a friction absorber 62A, which includesa rotor element or drum 64, and a stator element 66, carrying movablebrake shoes or friction pad elements 68. The drum 64 comprises a hub 96having a socket 98, in which is keyed an enlarged end 82 of the shaft 4.A circular plate 108 and a gasket 109 are secured to a flange 106 on thedrum 64 by cap screws 110.

The stator 66 comprises a cylindrical shaft portion 128, which projectsthrough a boss 114 on a circular plate 130 of substantially smallerdiameter than the inner diameter of the drum 64. The shaft 128 isrotatably mounted in the boss 114 and is also rotatably mounted in theboss 120 on the casing 42.

The friction pad means 68 comprises a pair of conventional arcuate brakeshoes 158 and 160, each having brake lining material 162 secured theretofor frictionally engaging the inner cylindrical surface of the drum 64when said shoes are moved outwardly, as will be readily understood. Theplate 130 carries a pair of adjusting pins 154 upon which the brakeshoes 158 and 160 are pivotally mounted.

Mounted on the plate 130 between the upper ends of the brake shoes 158and 160 is a conventional fluid pressure operated brake actuator unit176, which includes a cylinder 178 having a pair of piston-operated rods180 and 182 extending from the opposite ends thereof, it beingunderstood that the rods 180 and 182 are moved outwardly by fluidpressure within the cylinder 178 to expand the brake shoes 158 and 160.When fluid pressure is relieved in the cylinder 178, a return spring 174functions to retract the brake shoes 158 and 160 out of engagement withthe drum 64.

The shaft 128, FIG. 5, has an axial bore 184 which communicates at itsinner end with an axial bore 185 and a radial bore 186 in the plate 130.The radial bore 186 communicates with the interior of the cylinder 178for conducting fluid pressure to and from said cylinder. The outer endof the bore 184 is threaded to receive a fitting 188 to which a conduit246 is connected. Thus, by supplying fluid under pressure throughtheconduit 246 the actuator unit 176 can be operated to move the frictionpad means 68 into frictional engagement with the drum 64. The force withwhich such engagement is made will control the value of the resultantretarding force when the rotor or drum 64 is revolved. The magnitude ofsuch force can be controlled by varying the value of the fluid pressureapplied through conduit 246.

Mounted on the rear wall 44 of the coolant casing 42, FIGS. 4 and 5, andspaced from the shaft 128, is a transducer 192 for converting angularmovement of the stator 66 into fluid pressure, said transducer includinga housing 193 containing oil. An upwardly projecting push rod 194extends from a piston 195 mounted in the housing 193. A conduit 196 fullof oil is connected to the transducer 192, and leads to a fluid pressureoperated gauge 198. The transducer 192 is constructed so that when thepush rod 194 is depressed by mechanical force, hydraulic pressure willbe produced by the piston 195 within the transducer and will betransmitted to the gauge 198 through the conduit 196. The hydraulicpressure will be relieved when the mechanical force applied to the pushrod 194 is discontinued.

On the outer end of the shaft 128 is an arm 200 having a boss 202 fixedthereto by a key 204. The outer end 206 of the arm 200 rests on the pushrod 194 of the transducer 192. When the brake actuator unit 176 isoperated to move the friction pad means 68 into frictional engagementwith the drum wall 104 and the rotor 64 is rotated counter-clockwise (asviewed in FIG. 4), the shaft 128 and the stator 66 will tend to rotatetherewith. Angular movement of the stator 66 will engage the outer end206 of the arm 200 with the push rod 194, producing a fluid pressuresignal for transmission to the gauge 198. The value of the signal willbe proportional to the effective retarding force of the frictionabsorber 62A and, hence, the gauge 198 can be calibrated accordingly infoot pounds.

When the rotor 64 is rotated while the friction pad means 68 is inengagement with the drum wall 104, heat will be generated between therelatively stationary friction brake shoes 162 and the moving surface ofthe drum wall 104. This is dissipated by passing coolant through thecasing 42.

The load control apparatus for the friction absorber 62A includes theconduit 246 connected at one end to the fitting 188 for supplying fluidpressure to the brake actuator unit 176, the other end of the conduit246 being connected to an air-pressure-hydraulic-pressure transducer248, shown in corss-section in FIG. 2. The transducer 248 includes anupper housing section 250 and a lower housing section 252 having flanges254 and 256, respectively, on their confronting ends and between whichthe outer margin of a flexible rolling diaphragm 258 is clamped. Thehousing section 250 has a hollow lower portion 260, and a reduced hollowupper portion 262, the latter terminating in a boss 264 to which theconduit 246 is connected.

Received within the housing sections 250 and 252 is a member 266 havinga lower piston 268, and a relatively reduced upper plunger 270 slidablyreceived within the portion 262 of the housing section 250. The plunger270 carries a seal 272 in a groove 274 near its upper end. The lowerface 276 of the piston 268 is engaged with the diaphragm 258, and thelatter is secured thereto by a bolt 278 and a washer 280. The conduit246 and the chamber in the housing portion 262 above the plunger 270 arefilled with a suitable hydraulic fluid 282. Thus, when the member 266 ismoved upwardly, the fluid 282 will be pressurized by the plunger 270 foroperating the brake actuator unit 176.

The chamber in the lower housing section 252 has a port 284communicating therewith, to which is connected one end of a conduit 286leading from an air pressure control valve 288 connected to an airpressure source 290. When air pressure is supplied to the housingsection 252 beneath the rolling diaphragm 258, the piston 266 will bemoved upwardly to exert force on the fluid 282. The area 276 of thepiston 266 against which air pressure acts through diaphragm 258 isseveral times greater than the area of the upper end face 292 of theplunger 270, so that the pressure on the surface 276 will becorrespondingly multiplied in the fluid 282.

The control valve 288 is operable mechanically and automatically tocontrol the value of air pressure transmitted from the source 290 to thelower section 252 of the transducer 248. The valve 288 includes rightand left housing sections 294 and 296 (as viewed in FIG. 2) between theconfronting ends of which a rolling diaphragm 298 is clamped. Thehousing section 296 has a valve chamber 300 extending from the end face302 thereof, said valve chamber including a frustoconical seat 304 atits bottom, and terminating in a passage 306 leading to a larger chamber308. The chamber 308 has a frusto-conical side wall portion 310, whichterminates at a shoulder 312, and faces the diaphragm 298. A circularvalve seat 314 is secured to the shoulder 312. An inlet port 318communicates with the central portion of the valve chamber 300, and oneend of a conduit 320 extending from the pressure source 290 is connectedthereto. An outlet chamber 322 leads from chamber 308 and the region ofthe passage 306 to an outlet port 324, to which one end of the conduit286 is connected.

Received within the valve chamber 300 is a valve 326 having an enlargedhead 328 with a hemi-spherical surface engageable with the seat 304 toclose the passage 306, and a stem 330 which extends through the passage306 and through a central opening in the seat 314. The outer end of thevalve chamber 300 is closed by a plug 334 held in position by a plate336 secured to the housing section 296 by screws 338. A spring 340 iscompressed between the head 328 of the valve 326 and the plug 334, andfunctions to urge the spherical surface on said head into seatingengagement with the seat 304.

The diaphragm 298 has a central opening therein, through which projectsthe threaded end of a flanged member 342 having an axial passage 344,one end of which is frusto-conical to provide a seat 345 for receivingthe tip of the stem 330. The size of the passage 344 is chosen so thatwhen the tip of the stern 330 contacts the seat 345, said passage willbe closed. A diaphragm support member 348 having a central boss isthreaded on the member 342 and secures it to the diaphragm 298.

A cup-shaped member 352 is fitted over the boss on the member 348 andhas a wall 354 spaced from the end face of the members 342 and 348. Aplurality of circumferentially spaced passages 356 extend through thewall 354. The housing section 294 has a vent port 360 in the wallthereof, whereby air under pressure flowing through the passage 344 willtravel through the passages 356, and will exhaust through the vent port360.

The end wall 366 of the housing section 294 has a boss 368 into which aflanged guide 372 is threaded. The guide 372 has an axial bore 374through which a push rod 364 extends. The member 352 has a boss on theend thereof within which a socket is provided for the adjacent end ofthe push rod 364. The push rod 364 is actuated by a mechanical linkageunder the control of a speed responsive device driven by a belt 376 fromthe pulley 220. The speed responsive device includes a flyweightcentrifugal governor unit 378, comprising a cylindrical housing 380having a reduced extension 382 at one end within which a pair of spacedball bearings 384 is mounted. A shaft 386 rotates in the bearings 384,and has a pulley 388 secured to its outer end to be driven by the belt376. Thus, the shaft 386 will be rotated at a speed directlyproportional to the speed of the power input shaft 4.

The inner end of the shaft 386 has a cross member 390 from the ends ofwhich extend axial supports 392. One end of a fly-weight 394 is pivotedto each support 392 by a pin 396, so that when the shaft 386 is rotated,

the free ends of said fly-weights will be moved outwardly by centrifugalforce, as will be readily understood. Each of the weights 394 carries alug 398 which will move axially away from the shaft 386 as the weights394 swing outwardly upon rotation of said shaft.

The open end of the housing 380 is closed by a plate 400 having acentral boss 402 containing a bore 404. A cylindrical bushing 406 isfixed in the bore 404, and slidably receives the stem 408 of an actuatorelement 410 having a head 412 adjacent to which is attached a ballthrust bearing 414. The lugs 398 engage one race of the bearing 414 inorder to allow relative rotation between said lugs and the element 410.

When the shaft 386 is rotated and the weights 394 swing outwardly, thelugs 398 function to push the element 410 axially forwardly out of thehousing 308. Thus, the centrifugal unit 378 functions to convert rotarymovement of the shaft 386 into linear movement of the stem 408. The flyweights 394 and their lugs 398 are designed so that the element 410 willbe moved axially in direct relationship to centrifugal force acting onthe weights 394. Further, it is known that centrifugal force changes asthe square of the rotational speed. Thus, the element 410 will beshifted axially according to the square of the changes in speed of theshaft 386.

Movement of the element 410 is transmitted to the push rod 364 through aparallel lever arrangement 416, including a first lever 418, pivotallymounted at its upper end on a fixed pin 420 positioned in the same planeas the element 410. A second lever 422 extending parallel to the firstlever 418, and pivotally mounted at its lower end on a fixed pin 424 isalso disposed in the same plane as the push rod 364. The lower end ofthe lever 418 is engaged with the outer end of the push rod 364, and theupper end of the lever 422 is engaged by the outer end of the element410.

The levers 418 and 422 are spaced apart, and received there between isan adjustable fulcrum wheel 426 having guide flanges 428 for retainingthe levers 418 and 422 engaged with its outer surface. The fulcrum wheel426 is carried by a yoke 430, the lower end of which is connected by aconventional universal joint 432 to a swivel head 434.

The swivel head 434 has a bore 436 in the underface thereof, withinwhich a ball bearing 438 is secured by a snap ring 440. The reducedupper end 442 of a threaded rod 444 passes through the bearing 438 andis rotatably secured thereto by a snap ring 446. The rod 444 alsoextends through a threaded opening in a fixed plate 448, and has a handwheel 450 mounted on its lower end. Thus, by turning the hand wheel 450the fulcrum wheel 426 can be adjusted along the levers 418 and 422 froma position opposite the upper pivot pin 420 to a position opposite thelower pivot pin 424.

Depending upon where the fulcrum wheel 426 is positioned, the leverarrangement 416 can multiply or divide the overall axial movement of theelement 410 and effect a proportionate movement of the push rod 364, andwhich movement in any event, is correlated to the speed of the shaft386.

For example, if the fulcrum wheel 426 is positioned directly oppositethe upper pivot pin 420, no force would be transmitted to the push rod364 by the element 410 because then no outward movement of said elementcould occur. The air control valve 288 would therefore remain closedduring rotation of shaft 386, and no retarding action would be producedby the power absorption unit 2.

If the adjustable fulcrum wheel 426 is positioned midway between thepivot pins 420 and 424, the movement transmitted to the push rod 364will be directly proportional to the force exerted on the lever 422 bythe element 410. This force is then converted to a pro portional airsignal by the air control valve 288, and is transmitted to the airpressure-oil pressure signal by the ratio of the area 276 of the piston268 exposed to air pressure, to the area 292 of the plunger 270 exposedto the hydraulic fluid 282, and produces fluid pressure for transmissionthrough the conduit 246 to the brake actuator unit 176 of the frictionabsorber 62A. As the speed of the prime mover, and hence of the rotor 64of the friction absorber 62A is changed, the signal pressure to thebrake actuator unit 176 will be changed with the square of the change inspeed. The result is, that the retarding force produced by the frictionabsorber 62A will have the characteristics of the curve D in FIG. 6, andthe power curve for the friction absorber will be as shown at H in FIG.7.

If the adjustable fulcrum wheel 426 is positioned closer to the lowerpivot pin 424 than to the upper pivot pin 420, the force or movementtransmitted to the push rod 364 is multiplied, resulting in a higher airpressure being transmitted to the transducer 248, and in a greaterretarding force exerted by the power absorber unit 62A. Similarly, ifthe movable fulcrum wheel 426 is positioned nearer the upper pivot pin420 than the lower pivot pin 424, the signal pressure to the brakeactuator unit 176 will be less than when said fulcrum wheel ispositioned midway between the pivot pins 420 and 424. The force exertedon the movable friction pad elements 68 by the brake actuator unit 176can thus be easily varied anywhere between zero and maximum for anygiven speed of the input shaft 4, merely by adjusting the position ofthe fulcrum wheel 426.

In operation, when no inward pressure is exerted on the push rod 364,the head 328 of the valve 326 will be held in engagement with the seat304 by the coil spring 340, thus closing passage 306. No air pressurewill then flow from the conduit 320 into the conduit 286.

When the push rod 364 is moved inwardly, the head 328 will be disengagedfrom the seat 304 and an annular passage from the valve chamber 300 tothe passage 306 will be established. The size of this annular passage,and hence the rate of air flow through the passage 306 will vary withthe extent to which the valve 326 is opened, said valve being designedso that the change in area of said annular flow space will be directlyproportional to the inward movement of the push rod 364.

When the pressure on the push rod 364 is relieved, and said push rod isallowed to move outwardly, the spring 340 will return the valve 326 toseating engagement with the seat 304 for closing the passage 306.Thereafter, air pressure returned from the transducer 248 to the housingsection 296 through the conduit 286 will act on the diaphragm 298 tomove the same outwardly, thereby unseating the end of the valve stem 330from the seat 345 at one end of passage 344. Air pres sure will thenexhaust to atmosphere through the passage 344, the passages 356, and thevent port 360, whereby air pressure on the bottom face of the piston 268of the transducer 248 will be relieved. It is thus seen that bymanipulating the push rod 364, the pressure exerted by the brakeactuator unit 176 to engage the friction pad means 68 can be varied atwill, and that for each longitudinal position of the push rod 364 therewill be a corresponding air pressure established in the transducer 248,resulting in a corresponding fluid pressure being transmitted to thebrake actuator unit 176.

Referring now to the graph of FIG. 6, the driving torque-speedcharacteristics for a typical motor vehicle engine is indicated by thecurve A. Here, the values of torque in foot pounds are plotted asordinates, and the corresponding vehicle speeds in miles per hour areplotted as abscissas. It is seen that the engine driving torque risesrapidly with increasing engine speed, from O to about 25 miles per hour,and that thereafter driving torque increases at a slower rate withengine speed, until at about 47 miles per hour the driving torquebecomes stabilized. Over the range from about 47 miles per hour to about65 miles per hour, no appreciable increase in driving torque occurs.Above about 65 miles per hour, the value of engine driving torquedecreases with increasing engine speed.

One manner of operating the friction absorber 62A would be to supply aconstant fluid pressure to the brake actuator 176, the result of whichis illustrated by the curve B in FIG. 6, wherein it is seen that thevalue of the retarding force would then be constant over the completerange of engine speed, from 0 miles per hour upwardly.

While under-load testing a motor vehicle engine, or other prime mover,it is desirable to operate the engine at several different stableoperating speeds. When using friction dynamometer equipment, such astable speed is obtained by matching the value of the generatedretarding force to the value of the driving torque, until operation ofthe prime mover at the desired preselected speeds results. Turning tothe curves A and B in FIG. 6, it is seen that between about 47 and about65 miles per hour the engine driving torque curve A is parallel with theconstant value retarding force curve B. Because of this parallelrelationship, it is practically impossible over this common drivingspeed range to match retarding force to the driving torque and effectstability. The result is a hunting action, or a "'running wild of theengine, and hence true performance testing of the engine is notpossible.

Another problem, with a constant retarding force, results from the factthat in the lower speed ranges, driving torque decreases rapidly invalue with decreased speed, as is shown by the curve A, FIG. 6. Thus,assuming that under-load testing is in progress at a substantiallystable engine speed of 30 miles per hour, a problem arises if the engineshould misfire or momentarily lose power. While there would then be animmediate decrease in driving torque, the retarding force would remainconstant, and as the driving torque began to decrease, the constantretarding force would act to further slow the engine, and rapid decreasein driving torque would occur until the engine completely stalled. Thiscondition can be alleviated by having the retarding force increase anddecrease with changes in speed, and hence with driving torque.

It has been found that for the most efficient engine operation, thevalue of the retarding force should preferably be substantially zero atZero engine speed, and should rise and fall faster than the changes inthe driving torque output of the prime mover being tested. When thevalue of the retarding force is thus varied, the power absorbed, versusengine speed, will increase and decrease more rapidly than engine poweroutput.

Referring again to FIG. 6, curve C represents a situation whereretarding force is varied directly with changes in engine speed, whichcan be done by varying the value of the fluid pressure supplied to thebrake actuator unit 176 in direct proportion to changes in the speed ofthe engine being tested.

Thus, at zero speed the retarding force is also zero. As the enginespeed increases, there is a corresponding increase in driving torque,and similarly, when engine speed decreases the retarding force changesaccordingly. It is seen that the retarding force curve C cuts sharplyacross the typical driving torque curve A, at about 58 miles per hour,and that there are no regions where the retarding force curve C isparallel with the engine driving torque curve A. Thus, retarding forcecan easily be matched with driving torque to provide a stable operatingspeed, and there is no problem of engine stall occurring when there is atemporary decline in driving torque, because the retarding force followssuch decline.

In the case of automotive engines, it has been found that the bestrelationship for retarding force is to have the value thereof increaseand decrease as approximately the square of the change in speed, and tobe at zero speed. The reason this is a nearly ideal condition is that itvery closely simulates the load actually imposed on a conventionalautomobile engine while the vehicle is being driven on a level road.Such a retarding force versus speed curve is shown at D in FIG. 6. Suchretarding force can be created by varying the value of the pressure onthe friction pad elements 162 in accordance with the square of theengine speed.

The load control apparatus or system of the present invention will vary,with changes in speed, the pressure with which the friction pad elements162 are urged into frictional engagement with the drum wall 104 of therotor 64. The system, therefore, is effective to vary the value of theretarding force produced in proportion to the driving torque. The loadcontrol apparatus is designed so that the retarding force versus speedcurve of the friction absorber 62A can be shifted to the right or leftaround zero in FIG. 6 to obtain nearly any desired value of retardingforce at any given speed, whereby nearly any stable operating speed canbe established for under-load testing of a prime mover.

FIG. 7 is a graph wherein road horsepower is plotted against vehiclespeed in miles per hour, the curve E showing a typical power curve foran automobile engine. A typical power curve for a friction absorberwherein the retarding force is constant is shown at F, and it is seenthat the slope of the curve F is substantially less than that of thecurve E, whereby the power absorbed by the friction absorber 62A risesand falls at a slower rate with speed than does engine power. On theother hand, the curve H plots the power absorbed by the frictionabsorber 62A against engine speed in terms of vehicle speed in miles perhour, and it is seen that in this instance the absorbed power curve hasa slope substantially greater than the engine power curve E, whereby theabsorbed power rises and falls at a rate faster than the increases anddecreases in engine power. The arrangement of FIGS. 1 to 5 thus make itpossible to easily attain any desired stable operating speed for aprimemover during under-load testing, and because retarding force andabsorbed power rise and fall at faster rates than driving torque andengine power, rapid response of the friction absorber 62A to changes invehicle speed is assured and the problem of stalling in instances wherethe prime mover momentarily loses power is eliminated.

Under certain conditions it is possible to substitute a pneumaticallyoperated brake actuator unit in place of the hydraulic brake actuatorunit 176, FIG. 1, for moving the friction members 162 into engagementwith the rotor drum 64. If a pneumatic brake actuator unit issubstituted for the hydraulic actuator unit 176, then the air-pressurefluid-pressure transducer 248 can be eliminated, and the frictionelements 162 would then be operated directly by the air pressure outputfrom the valve 288.

Obviously, many additional modifications and variations of the presentinvention are possible in the light of the above teachings.

I claim:

1. Load control means for controlling the retarding forcecharacteristics of a rotary power absorber for absorbing driving torquefrom a shaft while said shaft is being driven from a prime mover, saidpower absorber including means for producing retarding force, and fluidpressure responsive means for causing said retarding means to produce aretarding force to oppose said driving torque, comprising: speedresponsive pressure regulating means including a governor to be drivenfrom the prime mover for providing a fluid pressure corresponding invalue to, and varying as, the speed of rotation of the prime mover andincluding an air pressure control valve connected with a source ofsupply of air under pressure, said air valve being actuated by saidspeed responsive means, and means including anairpressure-to-hydraulic-pressure transducer interposed between the airpressure control valve and the means for producing the retarding forcefor transmitting the controlled fluid pressure from said air pressurecontrol valve to said fluid pressure responsive means to apply aretarding force of a corresponding value to the power absorber.

2. Load control means for controlling the retarding forcecharacteristics of a rotary power absorber for absorbing driving torquefrom a shaft while said shaft is being driven from a prime mover, saidpower absorber including means for producing retarding force, and fluidpressure responsive means for causing said retarding means to produce aretarding force to oppose said driving torque, comprising: speedresponsive pressure regulating means including a governor to be drivenfrom the prime mover for providing a fluid pressure corresponding invalue to, and varying as, the speed of rotation of the prime mover andincluding an air pressure control valve connected with a source ofsupply of air under pressure, an adjustable linkage disposed betweensaid governor and said valve, actuated by said speed responsive means,and means for transmitting the controlled fluid pressure from said valveto said fluid pressure responsive means to apply a retarding force of acorresponding value to the power absorber.

3. Load control means as defined in claim 2, wherein the adjustablelinkage comprises a pair of generally parallel bars with a fulcrummember therebetween, one bar being pivotally mounted at one end and theother pivotally mounted at the end thereof remote from said one end, andwherein the governor includes a movable element acting on the free endof one of said bars, and the fluid pressure control valve includes amember acted upon by the free end of the other of said bars.

4. Load control means as defined in claim 3, wherein the fulcrum memberbetween the bars is a roller, and means is connected with said rollerfor adjusting its position along the length of said bars.

5. Load control means for controlling the retarding forcecharacteristics of a rotary power absorber for absorbing driving torquefrom a shaft while said shaft is being driven from a prime mover, saidpower absorber including means for producing retarding force, andhydraulic fluid pressure responsive means for causing said retardingmeans to produce a retarding force to oppose said driving torque,comprising: speed responsive pressure regulating means to be driven fromthe prime mover for providing a fluid pressure corresponding in valueto, and varying as, the speed of rotation of the prime mover andincluding fluid pressure control means actuated by said speed responsivemeans, said fluid pressure control means including a flow control valvehaving an inlet and an outlet, a source of air supply under pressureconnected to the flow control valve inlet, and means for transmittingthe controlled air pressure from the outlet of said flow control valveto said hydraulic fluid pressure responsive means to apply a retardingforce of a corresponding value to the power absorber and including anair-pressure-to-hydraulicpressure transducer for transforming thecontrolled air pressure to hydraulic pressure.

6. Load control means as defined in claim 5, wherein the speedresponsive pressure regulating means includes a governor device operablein response to variations in the speed of the prime mover; a linkageactuated by force supplied by said governor; and an adjustable fulcrumelement for increasing or decreasing the force developed by saidgovernor, the flow control valve being arranged to be actuated by saidlinkage in proportion to the force exerted thereon by said governor.

7. Load control means for controlling the retarding forcecharacteristics of a friction type rotary power absorber for absorbingdriving torque from a shaft while said shaft is being driven from aprime mover, said power absorber including friction brake means forproducing retarding force, and fluid pressure responsive means forcausing said friction brake means to produce a retarding force to opposesaid driving torque, comprising: speed responsive pressure regulatingmeans to be driven from the prime mover for providing a fluid pressurecorresponding in value to, and varying as the speed of rotation of, theprime mover and including a source of air pressure; a conduit connectedwith said source of air pressure; an air pressure control valveconnected with said conduit for controlling the air pressure from saidair source, said air pressure control yalve being operated in responseto he speed of rotation of the prime mover to control the fluid pressureto said fluid pressure responsive means to actuate said friction brakemeans, the value of said controlled air pressure being a pre-selectedfunction of the speed of rotation of the prime mover; and means fortransmitting the controlled fluid pressure from said air pressurecontrol valve to said fluid pressure responsive means to apply aretarding force of a corresponding value to the power absorber.

8. A friction type power absorber for use in analyzing the performanceof a prime mover, comprising: a driven brake shaft for receiving thedriving torque from a prime mover; power absorption means includingrotor means connected to receive driving torque from said shaft, statormeans operatively disposed relative to said rotor means; friction brakemeans carried by one of either said rotor or stator means and movableinto and out of frictional engagement: with the other; and fluidpressure operated actuator means operable to apply force for moving saidfriction brake means into said frictional engagement to thereby applyretarding force to said rotor in opposition to driving torque appliedthereto by said shaft, the value of said retarding force varying withthe force exerted by said actuator means; and control means includingmeans to generate a pressure signal to actuate an element of a fluidsystem for controlling said actuator means by fluid pressure regulatedin accordance with the speed of rotation of said shaft so that the forceexerted on said friction brake means by said actuator means varies as afunction of said rotational speed and so that said retarding force issubstantially zero at zero speed of said shaft and increases anddecreases in value at a rate faster than said driving torque.

9. A friction type power absorber as defined in claim 8, wherein thevalue of the retarding force is varied in accordance with a mathematicalfunction of the speed of rotation of the prime mover.

10. A friction type power absorber as defined in claim 8, wherein thevalue of the retarding force is varied in accordance with substantiallythe square of the speed of rotation of the prime mover.

11. A friction type power absorber as defined in claim 8, wherein thecontrol means includes a speed responsive governor.

12. A friction type power absorber as defined in claim 11, wherein thegovernor is driven at a speed proportional to the speed of the brakeshaft driven from the prime mover.

1. Load control means for controlling the retarding forcecharacteristics of a rotary power absorber for absorbing driving torquefrom a shaft while said shaft is being driven from a prime mover, saidpower absorber including means for producing retarding force, and fluidpressure responsive means for causing said retarding means to produce aretarding force to oppose said driving torque, comprising: speedresponsive pressure regulating means including a governor to be drivenfrom the prime mover for providing a fluid pressure corresponding invalue to, and varying as, the speed of rotation of the prime mover andincluding an air pressure control valve connected with a source ofsupply of air under pressure, said air valve being actuated by saidspeed responsive means, and means including anair-pressure-tohydraulic-pressure transducer interposed between the airpressure control valve and the means for producing the retarding forcefor transmitting the controlled fluid pressure from said air pressurecontrol valve to said fluid pressure responsive means to apply aretarding force of a corresponding value to the power absorber.
 2. Loadcontrol means for controlling the retarding force characteristics of arotary power absorber for absorbing driving torque from a shaft whilesaid shaft is being driven from a prime mover, said power absorberincluding means for producing retarding force, and fluid pressureresponsive means for causing said retarding means to produce a retardingforce to oppose said driving torque, comprising: speed responsivepressure regulating means including a governor to be driven from theprime mover for providing a fluid pressure corresponding in value to,and varying as, the speed of rotation of the prime mover and includingan air pressure control valve connected with a source of supply of airunder pressure, an adjustable linkage disposed between said governor andsaid valve, actuated by said speed responsive means, and means fortransmitting the controlled fluid pressure from said valve to said fluidpressure responsive means to apply a retarding force of a correspondingvalue to the power absorber.
 3. Load control means as defined in claim2, wherein the adjustable linkage comprises a pair of generally parallelbars with a fulcrum member therebetween, one bar being pivotally mountedat one end and the other pivotally mounted at the end thereof remotefrom said one end, and wherein the governor includes a movable elementacting on the free end of one of said bars, and the fluid pressurecontrol valve includes a member acted upon by the free end of the otherof said bars.
 4. Load control means as defined in claim 3, wherein thefulcrum member between the bars is a roller, and means is connected withsaid roller for adjusting its position along the length of said bars. 5.Load control means for controlling the retarding force characteristicsof a rotary power absorber for absorbing driving torque from a shaftwhile said shaft is being driven from a prime mover, said power absorberincluding means for producing retarding force, and hydraulic fluidpressure responsive means for causing said retarding means to produce aretarding force to oppose said driving torque, comprising: speedresponsive pressure regulating means to be driven from the prime moverfor providing a fluid pressure corresponding in value to, and varyingas, the speed of rotation of the prime mover and including fluidpressure control means actuated by said speed responsive means, saidfluid pressure control means incLuding a flow control valve having aninlet and an outlet, a source of air supply under pressure connected tothe flow control valve inlet, and means for transmitting the controlledair pressure from the outlet of said flow control valve to saidhydraulic fluid pressure responsive means to apply a retarding force ofa corresponding value to the power absorber and including anair-pressure-to-hydraulic-pressure transducer for transforming thecontrolled air pressure to hydraulic pressure.
 6. Load control means asdefined in claim 5, wherein the speed responsive pressure regulatingmeans includes a governor device operable in response to variations inthe speed of the prime mover; a linkage actuated by force supplied bysaid governor; and an adjustable fulcrum element for increasing ordecreasing the force developed by said governor, the flow control valvebeing arranged to be actuated by said linkage in proportion to the forceexerted thereon by said governor.
 7. Load control means for controllingthe retarding force characteristics of a friction type rotary powerabsorber for absorbing driving torque from a shaft while said shaft isbeing driven from a prime mover, said power absorber including frictionbrake means for producing retarding force, and fluid pressure responsivemeans for causing said friction brake means to produce a retarding forceto oppose said driving torque, comprising: speed responsive pressureregulating means to be driven from the prime mover for providing a fluidpressure corresponding in value to, and varying as the speed of rotationof, the prime mover and including a source of air pressure; a conduitconnected with said source of air pressure; an air pressure controlvalve connected with said conduit for controlling the air pressure fromsaid air source, said air pressure control valve being operated inresponse to the speed of rotation of the prime mover to control thefluid pressure to said fluid pressure responsive means to actuate saidfriction brake means, the value of said controlled air pressure being apre-selected function of the speed of rotation of the prime mover; andmeans for transmitting the controlled fluid pressure from said airpressure control valve to said fluid pressure responsive means to applya retarding force of a corresponding value to the power absorber.
 8. Afriction type power absorber for use in analyzing the performance of aprime mover, comprising: a driven brake shaft for receiving the drivingtorque from a prime mover; power absorption means including rotor meansconnected to receive driving torque from said shaft, stator meansoperatively disposed relative to said rotor means; friction brake meanscarried by one of either said rotor or stator means and movable into andout of frictional engagement with the other; and fluid pressure operatedactuator means operable to apply force for moving said friction brakemeans into said frictional engagement to thereby apply retarding forceto said rotor in opposition to driving torque applied thereto by saidshaft, the value of said retarding force varying with the force exertedby said actuator means; and control means including means to generate apressure signal to actuate an element of a fluid system for controllingsaid actuator means by fluid pressure regulated in accordance with thespeed of rotation of said shaft so that the force exerted on saidfriction brake means by said actuator means varies as a function of saidrotational speed and so that said retarding force is substantially zeroat zero speed of said shaft and increases and decreases in value at arate faster than said driving torque.
 9. A friction type power absorberas defined in claim 8, wherein the value of the retarding force isvaried in accordance with a mathematical function of the speed ofrotation of the prime mover.
 10. A friction type power absorber asdefined in claim 8, wherein the value of the retarding force is variedin accordance with substantially the square of the spEed of rotation ofthe prime mover.
 11. A friction type power absorber as defined in claim8, wherein the control means includes a speed responsive governor.
 12. Afriction type power absorber as defined in claim 11, wherein thegovernor is driven at a speed proportional to the speed of the brakeshaft driven from the prime mover.